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Contract No. DE-AC36-08GO28308

Advanced Liquid Cooling for a 

Traction Drive Inverter Using Jet 

Impingement and Microfinned 

Enhanced Surfaces

Preprint

S.K. Waye, S. Narumanchi, M. Mihalic, 

G. Moreno, K. Bennion, and J. Jeffers

Presented at ITherm 2014

Orlando, Florida

May 27 – 30, 2014

Conference Paper

NREL/CP-5400-61060

August 2014
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Advanced Liquid Cooling for a Traction Drive Inverter Using 

Jet Impingement and Microfinned Enhanced Surfaces

Scot K. Waye, Sreekant Narumanchi, Mark Mihalic, Gilbert Moreno, Kevin Bennion, Jana Jeffers

National Renewable Energy Laboratory

15013 Denver West Parkway

Golden, Colorado 80401

Email: scot.waye@nrel.gov

ABSTRACT

This study evaluates a jet impingement based cooling 

strategy combined with microfinned enhanced surfaces as a 

means of improving thermal management for power electronic 

devices. For comparison, a baseline channel flow heat 

exchanger and jet impingement on plain surfaces are 

characterized. The jets, augmented with enhanced microfinned 

surfaces, provide localized cooling to areas heated by the 

insulated-gate bipolar transistors and diode devices. Lighter 

materials and simpler manufacturing while managing required 

pumping power increase the overall performance while 

reducing weight, volume, and cost. Computational fluid 

dynamics modeling validated by experiments was used to 

characterize the baseline as well as jet-impingement-based 

heat exchangers at typical automotive flow rates using a 50%–

50% mixture by volume of water and ethylene glycol. The 

three cooling configurations were tested at full inverter power 

(40 to 100 kW output power) on a dynamometer. An increased 

thermal performance was observed for the jet-impingement 

configurations. Experiments were also performed to 

investigate the long-term reliability of the jets impinging on

enhanced surfaces.

KEY WORDS: jet-impingement, enhanced surfaces, 

microfinned surfaces, single-phase liquid cooling, power 

electronics, inverter thermal management, heat transfer

NOMENCLATURE

CFD computational fluid dynamics

DBA direct-bond-aluminum

DBC direct-bond-copper

IGBT insulated gate bipolar transistor

P pressure, Pa

Q heat, W

R resistance, K/W

T temperature, K or oC

V volumetric flow rate

WEG water-ethylene glycol coolant

Subscripts

avg average

hx heat exchanger

j junction

j-l junction to liquid coolant

l liquid coolant

max maximum

th thermal

INTRODUCTION

With the increasing trend towards miniaturization and 

advancements in devices, thermal management of power 

electronics is becoming more challenging. With more power 

passing through smaller areas, the heat flux increases and 

necessitates more aggressive cooling to maintain device 

operational temperatures. Reducing the component costs and 

allowing for increased power density and specific power are 

key objectives of the U.S. Department of Energy’s Advanced 

Power Electronics and Electric Motors Program. The program

targets are intended to increase market penetration of hybridelectric and all-electric vehicles.

The objective of this work is to design, develop, 

characterize, and demonstrate a light-weight, low-cost,

inverter-scale (based on a commercially available inverter),

single-phase liquid-cooled heat exchanger that is a significant 

improvement over conventional channel-flow-based heat 

exchangers. The purpose is to use a conventional coolant 

(water-ethylene glycol [WEG]) to maintain existing cooling 

architecture while improving thermal management of the 

inverter.

In previous work, we developed a first prototype [1] that 

showed promising thermal improvements using a submerged 

liquid jet-impingement array, but with an increase in the 

pumping power. A second prototype was designed, fabricated,

and tested. The results from this second prototype will be 

compared to the first jet-based prototype as well as the 

baseline channel flow configuration. 

General studies of single-phase liquid jets are extensive in 

the literature [2-5], including experimental, theoretical, and 

numerical approaches. Different configurations of impinging 

jets have been studied, including single free-surface jets [6], 

multiple free-surface jets [7-10], single submerged jets [6, 11, 

12], multiple submerged jets [7, 10, 11, 13], and confined 

single submerged jets [4, 14-16]. Both planar and circular jets 

have been studied.

For power electronics cooling applications, a large number 

of experimental correlations have been developed for the local 

and average heat transfer coefficients for impinging jets on the 

surface of a simulated chip [3-5], usually 10-mm by 10-mm or 

12.7–mm by 12.7-mm in area. Further parametric and 

optimization study examined potential improvements [17].

Ndao et al. [18] compare single and multiple submerged jetimpingement cooling strategies for electronics cooling with 

other technologies such as parallel micro-channel, in-line and 

staggered circular pin-fin, and offset strip-fin heat exchangers. 

Jet impingement combined with a mini-channel heat 

exchanger or flows through narrow gaps has also been 

investigated [19, 20]. Skuriat and Johnson [21] discuss direct 
2

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cooling on an aluminum nitride substrate with two insulated 

gate bipolar transistors (IGBTs) and two diode dies. Of high 

interest is cooling of wide band-gap devices, including gallium 

nitride devices, because maximum junction temperatures can 

be higher than silicon devices, and this was recently examined

by Bhunia et al. [22].

Air jets have also been studied extensively [11]. Some of 

the non-dimensional heat transfer correlations developed from 

studies on air jets [11] can also be applied to liquid jets. 

Researchers have explored the impact of a vast array of 

parameters such as jet velocity, jet diameter, impact angle, 

nozzle-to-chip spacing, nozzle-to-nozzle spacing, turbulence 

levels, nozzle shapes, nozzle length, jet pulsations, jet 

confinement, chip-surface enhancement, and fluid properties 

on the chip-surface heat transfer coefficients. All these are 

covered in detail in comprehensive reviews and studies on 

liquid and air jets [2, 3, 4, 11, 23].

Moreover, studies have demonstrated that jet impingement 

heat transfer can be further enhanced through the use of 

surface roughening techniques for both liquid and air jets [24-

27]. Moreno et al. [28] showed superior performance of 

submerged jets impinging on a microfinned enhanced surface.

These results were noteworthy because submerged/flooded jet 

configuration is a more practical application than free jets [29, 

30].

In automotive power electronics, it is important to reduce 

cost, increase power density and specific power to help make 

electric-drive vehicles more cost-effective and increase their 

market acceptance and penetration. Traditionally, channel 

flow-based cold plates using a WEG mixture (50%–50% by 

volume) are used in a majority of hybrid electric vehicles for 

cooling power electronics components [29].

This work focuses on using existing liquid coolant (WEG) 

loops with improved cooling technologies on the inverter 

scale. Through experiments and numerical modeling, we 

investigate submerged liquid jet impingement on plain and 

enhanced surfaces. By modeling the system and validating the 

model with low-power experiments, we predict the thermal 

performance of a full inverter-scale system. We also test the 

inverter at full power conditions using a dynamometer with 

full switching controls. We conduct a few fundamental 

reliability tests in order to investigate potential concerns of 

clogged jets and surface degradation over time.

APPROACH

Design and Fabrication

Building on lessons learned from the first prototype 

fabrication and testing [1], a second prototype heat exchanger 

utilizing a submerged jet-impingement cooling approach and a 

light-weight plastic material was fabricated. The main 

differences of this prototype from the first are the material and 

manufacturing process and the flow path. The first prototype 

was fabricated using a selective-laser-sintering rapid 

prototyping process using automotive-grade glass-fiberreinforced nylon plastic. The second prototype was machined 

from Delrin, a high-temperature plastic, although this 

component could be injection molded from a variety of hightemperature plastics. 

The first prototype replaced the channel flow configuration 

with jets that impinge on the module surfaces. The flow path 

remained similar to that of the baseline design. The jets 

increased the overall pressure drop, so the second prototype, 

shown in Fig. 1, removed unnecessary flow path to reduce the 

pressure drop back to levels similar to the channel-flow 

baseline. The first prototype used screwed in nozzles while the 

second prototype implemented plates with drilled holes

serving as nozzles. The nozzles are 1.4 mm in diameter. Note 

eight nozzles in each nozzle plate. Each power module 

contains 8 diodes and 8 IGBTs. Each nozzle is located nearly 

under the center of a diode or IGBT. Parts of the manifold 

have been omitted from fabrication for ease of testing. The jet 

impingement cooling approach eliminates the use of thermal 

interface material between the base plate and the heat 

sink/cold plate and therefore is expected to improve thermal 

performance. Further details on the previous design are 

discussed by Narumanchi [31].

Fig. 1 Second light-weight, inverter-scale plastic heat 

exchanger prototype. Schematic of one of three power 

modules overlaid with location of thermocouples

(embedded in base plate) for full inverter testing. 

Experiments

Three half bridge power modules, fabricated by PowerEx, 

were used in all testing. 

Low power testing. Experiments were conducted to 

characterize the performance of the new heat exchanger (Fig. 

2). In these tests, four diodes on the center module were 

powered with approximately 105W.

Previous experiments characterized the first prototype as 

well as the baseline channel-flow-based aluminum heat 

exchanger [31]. The tests were carried out using WEG (50%–

50% mixture by volume) as the coolant at a temperature of 

70°C. A typical industry flow rate of 1.67x10-4 m3

/s (10 

L/min) as well as 3.34x10-5

, 8.35x10-5, and 1.34x10-4 m3/s (2,

5, and 8 L/min) were examined for the various heat 

exchangers. Using 70oC WEG (50%–50% by volume), these 

flow rates correspond to jet Reynolds number values based on 

nozzle diameter of approximately 500, 1,250, 2,000, and 2,500 

for the four flow rates. The average temperature of the diodes 

was measured using a transient thermal tester. Based on the 

information of diode junction temperature as well as the total 

power dissipated in the diodes, the total thermal resistance was 

computed. 

Inlet

Flow path

Nozzle

Electrical 

connections

Thermocouple 

locations

Outlet
3

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Fig. 2 Second prototype testing on the WEG loop.

Fig. 3 Wolverine Tube MicroCool finned surface technology 

on module (top left) and close-up (top right and 

bottom).

Two sets of modules were used on the second prototype 

testing. The first set had standard plain copper base plates. The 

second set had microfinned surface enhancement located 

under each IGBT/diode pair, as seen in Fig. 3, with a close-up 

of the microfinned surface. The microfinned surface was 

created by skiving copper base plate, neither adding nor 

removing material. The increase in surface area is estimated to 

be approximately eleven times the plain surface. The 

microfins were approximately 100 μm wide and added 

approximately 200 μm to the base plate (the space between 

fins is cut into the original surface).

Full inverter testing. Full-scale inverter testing was 

conducted on a dynamometer at UQM Technologies Inc. 

Three configurations were run: channel flow with plain 

surfaces, jet impingement with plain surfaces, and jet 

impingement with microfinned enhanced surfaces, as seen in 

Fig. 4. The tests were run at power levels of 40, 60, 80, and 

100 kW inverter power at set points of 360 V and 2,000 rpm. 

A 40%–60% water-ethylene glycol (by volume) coolant 

flowing at 30oC and 10 L/min was used. 

Fig. 4 Full power inverter testing on dynamometer (channel 

flow with plain modules configuration shown).

Two k-type thermocouples on each module were inserted 

into the mid-point of the copper base plate nearly centered 

under the IGBT/diode pair, as shown in Fig. 5 and Fig. 6, and 

centered above the microfinned surface when applicable. Note 

the difference in locations of the thermocouples relative to the 

cooling path. For the channel flow configuration, the 

thermocouples are not located near aluminum wall-fluid 

interfaces. For the jet configuration, the thermocouples are 

located nearly between two jets, each centered on the device 

above. At equilibrium, these thermocouple temperatures, the 

inlet and outlet coolant temperatures, and ambient temperature 

were recorded. 

The average coolant temperature (inlet to outlet) was 

subtracted from the average thermocouple temperature to yield

a temperature difference to compare thermal performance. 

Although an estimate of the heat dissipated by the devices 

could be calculated using the fluid thermal properties and 

temperature rise in the coolant, it is not very accurate as there 

could be other heat dissipation pathways in the inverter. We

assumed that the heat dissipated by the devices at each power 

level is the same for each configuration since the comparison 

between configurations is the main metric being examined.
4

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Fig. 5 Thermocouple locations of the channel flow 

configuration. View is from bottom (module base plate 

facing towards reader) to highlight fluid path (shaded).

Fig. 6 Thermocouple locations for the jet-impingement 

configurations (three modules shown with eight IGBTs 

and eight diodes each). Note the circles near center of 

IGBTs and diodes are the nozzle locations for the jets.

Flow enters from the left in two parallel paths (see Fig. 

1 for flow path).

Modeling

A computational fluid dynamics CFD model was created 

using the same geometry and properties as the experimental 

setups to predict the improvement in thermal and overall 

performance of the impinging jet designs over the baseline at a 

higher power level (2.5 kW). Using the experimental data 

obtained with four diodes powered to produce approximately 

105 W of heat, the thermal resistance was matched in the 

model by adjusting the solder layer resistance for the jetimpingement cases, maintaining that resistance (since the 

same modules were used for the jet-impingement and channel 

flow experiments), and adjusting the thermal grease layer 

resistance for the channel flow cold-plate case. The thickness 

and resistance of the solder and grease layers were previously 

unknown and slightly vary for each module and setup.

A k-ω turbulence model was used as it has been shown to 

yield good agreement with experimental jet impingement 

results [29, 32]. For consistency, this model was used for the 

channel flow and jet impingement configurations. The grid 

was adapted to yield a y+ (non-dimensional distance of grid 

centroids) of less than one on the impinging surface, and the 

average and maximum device temperatures were monitored. 

(All residuals were set to 10-3 except for energy, which was set 

to 10-6

). Up to 20 million cells were used in the simulations, 

using parallel processing up to 32 cores. A grid independence 

study was conducted and the pressure drop and temperatures 

are expected to be mesh-independent to within 5%. Similar 

modeling was conducted in [1] and [31].

After validating the model at low power (105 W from four 

diodes), all of the 24 IGBTs and 24 diodes were powered at a 

two-to-one ratio, yielding a heat generation rate of 1.944x109

W/m3 (70 W per device) for the IGBTs and 2.12x109 W/m3

(35 W per device) for the diodes, yielding 2,520 W of total 

heat dissipation.

A flow rate of 1.67x10-4 m3

/s (10 L/min) for WEG at 70°C 

was used for the simulations. Modeling microfins for the 

current study was excessively expensive. To produce a similar

thermal performance for the microfinned enhanced surfaces, 

the flow rate was doubled to double the jet velocity. Previous 

research [28] found that for the microfinned surface 

enhancement used, the heat transfer coefficient for a jet 

velocity with the enhanced surface was approximately the 

same as the coefficient at twice the jet velocity on a plain 

surface.

Reliability

Reliability tests are currently being conducted using a 65°C 

automotive-grade WEG (with corrosion inhibitors) on two 

nickel-plated microfinned surfaces at 5 m/s. A schematic of 

the test apparatus is shown in Fig. 7. Near-continuous jet 

impingement testing has been performed for several months, 

with a full year of testing planned. The microfinned enhanced 

surface samples are characterized using digital microscopy for 

surface images, and the heat transfer coefficients are 

measured.

Fig. 7 Test setup for long-term reliability tests.
5

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Table 1. Thermal resistance based on four powered diodes, Rth,j-l [K/W], and improvement over channel flow heat exchanger for 

8 L/min and 10 L/min: (1) designates first prototype, (2) designates the second prototype.

RESULTS AND DISCUSSION

Low Power Experimentation

Experimentation using a transient thermal tester (T3ster) 

yields structure functions that provide information on the 

thermal capacitance and the thermal resistance. This 

information can be used to differentiate the thermal resistance 

of the passive stack and the cooling technology. By mapping 

all the resistances for various flow rates and cooling 

technologies (channel flow cold plate, impingement on plain 

surfaces, impingement on enhanced surfaces), the point where 

they diverge indicates the boundary between the passive stack 

in the module and the convective cooling technology. A map 

of thermal resistance for various heat exchanger 

configurations at 1.67x10-4 m3

/s (10 L/min) are given in Fig. 

8. In the case of the channel flow, the cold plate layer is 

included in the convective cooling technology as it is omitted 

for the impinging jet configurations. For the modules used in 

this testing, the thermal resistance of the passive stack was 

approximately 0.09 K/W under the condition of four diodes 

being powered.

Fig. 8 Thermal resistance map from T3ster for thermal 

response of four powered diodes.

For 1.34x10-4 m3

/s and 1.67x10-4 m3/s flow rates (8 L/min 

and 10 L/min), the junction to liquid thermal resistances (Rth,j-l

[K/W]) and percent reduction as compared to the baseline 

channel flow cold plate are provided in Table 1. Uncertainty in 

the resistance values were ±0.001 K/W and in the percent 

reduction were ±1%. The test cases included the channel flow, 

the first prototype jet-impingement heat exchanger with plain 

surfaces, and the second prototype jet-impingement heat 

exchanger with plain and microfinned enhanced surfaces. 

The thermal performance for the first and second 

prototypes with jets impinging on plain surfaces was nearly 

the same, which was expected because the jet design was 

essentially the same for both cases. For the jet impingement on 

a plain surface, the thermal resistance was reduced by 5.1% as 

compared with the performance of the baseline channel flow 

heat exchanger. On the microfinned surface, the reduction was 

10.3% at the nominally designed target distance (0.5 mm).

The resistance of the passive stack was estimated to be 

approximately 0.09 K/W. By subtracting this resistance from 

the total resistance, the convective resistance (the resistance 

from the solid-to-fluid), Rth,hx [K/W], was 15.2% lower for jet 

impingement on plain surfaces, and 37.0% lower for jet

impingement on enhanced surfaces, both with respect to the 

baseline system. Additional reduction to the junction-to-liquid 

resistance can be achieved by more aggressive cooling 

strategies or reducing the resistance in the passive stack.

For the jet-impingement heat exchangers, the nozzle-tobase plate distance (target distance) was also examined. The 

nominal target distance was 0.5 mm. Offsets of 1.45 mm and 

3.25 mm were introduced to examine the effect of increased 

distance between the nozzle and surface. For the microfinned

enhanced surface, the microfins extended approximately 200 

μm into the flow, reducing the jet length to 1.75 mm and 3.55 

mm with the two offsets, respectively.

Increased distance between the nozzle and the target 

reduces the submerged jet momentum. The interaction of the 

jet with the surface to create a thin boundary layer and 

turbulence are important aspects to transfer the heat from the 

solid to the fluid. For very low velocities (flow rate) and large 

target distances, the momentum of the jet decreases and the jet 

core spreads, decreasing heat transfer on the surface. For the 

plain surface, this leads to reduced heat transfer. For the 

microfinned surface, the interaction between the jet and the 

increased surface area over which it impinges is slightly 

increased, which helps increase the heat transfer. This effect is 

reduced with further distance since the jet is weakened further. 

For enhanced surfaces, there is an optimal distance depending 

on the jet characteristics.

Table 2 shows the experimental pressure drop across the 

heat exchanger configurations at 1.34x10-4 and 1.67x10-4 m3

/s

(8 and 10 L/min). As previously discussed, the first prototype 

had a large increase in pressure loss due to the introduction of 

the jets. By removing unnecessary flow path in the second 

prototype, the pressure loss was closer to the baseline channel 

Channel Flow Jet/Plain (1) Jet/Plain (2) Jet/Microfinned (2)

Target distance (mm) N/A 0.5 0.5 1.95 0.3 1.75 3.55

Resistance, Rth,j-l [K/W]

1.34x10-4 m3/s (8 L/min) 0.139 0.136 0.143 0.126 0.126 0.128

1.67x10-4 m3/s (10 L/min) 0.136 0.128 0.129 0.136 0.122 0.119 0.122

Improvement over baseline

1.34x10-4 m3/s (8 L/min) 2.2% -2.9% 9.4% 9.4% 7.9%

1.67x10-4 m3/s (10 L/min) 5.9% 5.1% 0.0% 10.3% 12.5% 10.3%
6 

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flow heat exchanger. Therefore, the improvement in the 

coefficient of performance, defined as the inverse product of 

the resistance and pumping power (1/RthΔPV), was nearly the 

same for the jet impingement on the plain surface heat 

exchanger and 13% better for the jet impingement on the 

microfinned surfaces heat exchanger. 

Table 2. Experimental pressure drop [Pa] across heat 

exchanger. (1) denotes first prototype, (2) denotes second 

prototype. Nozzle diameter 1.4 mm (48 nozzles). 

Flow Rate [m3

/s] (L/min) 1.34x10-4 (8) 1.67x10-4 (10) 

Channel Flow 12,686 19,598 

Jet/Plain (1) 24,407 

Jet/Plain (2) 13,221 20,581 

Jet/Microfinned (2) 12,755 19,995 

Heat Exchanger Modeling 

The focus of the modeling was for an IGBT-to-diode 

power ratio of two to one and a total heat generation of 2,520 

W. Flow rates, pressure drops, average and maximum 

temperatures, and junction to liquid thermal resistances based 

on the respective temperatures are given in Table 3. 

The thermal resistance for the jet-impingement heat 

exchanger was reduced by 9% and 32% for the plain and 

microfinned surfaces, respectively. These improvements 

translated to reductions in the average and maximum 

temperatures of approximately 5°C (Fig. 9) and 15°C for the 

plain and microfinned surfaces with jet impingement, 

respectively, as compared to the baseline heat exchanger. 

The reductions in thermal resistance from the inverter 

power modeling can be considered an idealized situation 

yielding the highest potential improvement in thermal 

performance. In the experiments, some heat from the devices 

goes into the module packaging and inverter case, so not all 

heat transfers to the coolant. Therefore, the thermal resistance 

is artificially reduced (Rth,j-l = ΔT/Q). Validation of the model 

assumed that all heat transferred to coolant. The geometry in 

the model included up to the device (no module casing), with 

boundary conditions on external faces being adiabatic. 

The localized and directed cooling from impinging jets 

creates high local heat transfer coefficients where the heat flux 

is the highest, under the devices, as shown in Fig. 10. 

The thermal resistance is subject to the convective cooling 

resistance, which relies on the cooling strategy or technology, 

and the heat flux magnitude and distribution. As the heat seeks 

the path of least resistance, either due to a low convective 

cooling resistance at the solid-fluid interface or the absence of 

multiple heat generating devices, it will manifest as a lower 

thermal resistance due to reduced heat spreading. The low 

power test only heated four diodes while the full model heats 

all diodes and IGBTs. 

Using basic analytical equations for device losses, the 

increase in potential power can be calculated from the 

increased heat dissipated due to the decreased thermal 

resistance. The specific power increased 29% and 55%, and 

power density increased 6% and 28% over the baseline for the 

plain and microfinned surfaces with jet impingement, 

respectively. The coefficient of performance improved by up 

to 5% and 40%, respectively. 

A 9% decrease in thermal resistance allows for 10% more 

heat to be dissipated. Using the same analytical equations 

mentioned before, this correlates to an approximate 6% 

increase in power per device area, or a 6% decrease in device 

area per power by varying only the current. Therefore, 6% of 

the device area (such as silicon) or number of devices could 

theoretically be removed, creating a cost savings. Similarly, a 

32% decrease in thermal resistance translates to an 

approximate device area reduction of 22%. 

Fig. 9 Temperatures of modules for baseline (top) and jet 

impingement on the plain surface for the first prototype 

(bottom). The second prototype has similar thermal 

performance. 

Fig. 10 Heat transfer coefficient at the solid-fluid interface for 

baseline (top) and jet impingement on the plain surface 

of the first prototype (bottom). The second prototype 

has similar thermal performance. 
7

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Table 3. Thermal resistance and device temperatures for modeled heat exchangers. (Rth,j-l = (Tdevices-Tcoolant)/Q, where Tcoolant is 

70oC [343.2 K] and Q is 2,520 W). *Flow rate doubled to match heat transfer coefficient of microfinned enhanced surface.

Full Inverter Experimentation

Information provided by the UQM Technologies Inc., who 

supplied the baseline inverter and conducted the dynamometer 

testing, estimated several outputs using a model that they have 

developed. For the inverter power levels (output shaft power) 

of 40, 60, 80, and 100 kW, the heat dissipated (losses) were 

estimated to be 0.9, 1.3, 1.6, and 2.2 kW. For comparison, by 

using the mass flow rate, the specific heat, and the inlet-outlet 

coolant temperature difference, the heat dissipated was 

estimated to be 0.9, 1.2, 1.7, and 1.9 kW, respectively, an

estimate expected to be low since all the heat is not transferred 

to the liquid coolant. The estimated junction temperatures of 

the IGBTs and diodes at the four power levels for the baseline 

inverter from their model are shown in Table 4. Recall that the 

coolant temperature for these tests was 30oC. The model that 

predicts these values has an estimated 15% uncertainty.

Table 4. Predicted IGBT and diode junction temperatures [oC].

Power [kW] 40 60 80 100

IGBT 46.6 54.5 62.2 73.2

Diode 47.7 52.6 56.3 64.1

Average 47.2 53.6 59.3 68.7

The average measured temperatures at the thermocouple 

locations are given in Table 5. The agreement with the model 

is relatively close, although the measured temperatures overall 

are lower than modeling results because the location of the 

thermocouples is in the base plate rather than at the device 

junction.

Table 5. Average thermocouple temperatures for three 

configurations [oC].

Power [kW] 40 60 80 100

Channel, Plain 46.6 52.5 60.2 68.7

Jet, Plain 53.5 60.0 67.6

Jet, Microfinned 44.8 50.2 56.1 62.3

The thermocouple locations caused a greater variation in 

temperatures for the channel flow configuration than the jet 

impingement configurations, as seen in Fig. 11. For the 

impinging jets, the temperatures (±0.8oC) for thermocouples 1, 

3, and 5 are generally slightly higher than thermocouples 2, 4, 

and 6, indicating that the flow through the flow paths with 

thermocouples 2, 4, and 6 is likely higher than for 1, 3, 5 

(higher flow rate equates to a higher jet velocity and heat 

transfer coefficient).

Fig. 11 Thermocouple (TC) temperatures for channel flow 

with plain modules and jet impingement with 

microfinned modules. 

The temperature reading for the channel flow configuration 

was more dependent on the location of the thermocouple 

relative to the channel flow path. As seen in Fig. 5, the 

thermocouples were located nearly above solid-fluid interfaces 

of the coolant flow paths. Each jet provides nearly the same jet 

velocity and coolant temperature, whereas the coolant in the 

channel flow varies along the parallel and serial paths. This 

temperature variation between the hottest and coldest 

thermocouple temperature measurement is shown in Table 6.

The jet configuration provides more temperature uniformity in 

the system than the channel flow configuration, thus cooling 

devices more equally.

Table 6. Temperature variation [oC] (isothermality) between 

hottest and coldest thermocouple temperatures.

Power [kW] 40 60 80 100

Channel, Plain 7.7 11.2 14.3 15.8

Jet, Plain 7.6 10.0 12.3

Jet, Microfinned 4.3 5.9 7.2 9.8

The local cooling afforded by the jets is one advantage to 

the technology, as the highest cooling is located near the 

highest heat flux and temperature. This provides for a more 

uniform temperature distribution among the various devices 

whereas for the channel flow technology, some devices are 

hotter than others. In the channel flow heat exchanger, the heat 

is spread by the additional cold plate layer and then is 

removed by the fluid. Thus, more conduction (and thermal 

Flow rate 

[m3/s] (L/min)

∆P

[Pa]

Tavg,devices

[K]

Rth,j-l

[K/W]

Tmax,devices

[K]

Rth,j-l

[K/W]

Channel Flow Cold Plate 1.67x10-4 (10) 17,891 391.6 0.0192 395.8 0.0209

Jet/Plain (1) 1.67x10-4 (10) 20,800 386.7 0.0173 390.6 0.0188

Jet/Microfinned (1) 3.34x10-4 (20)* 376.6 0.0132 380.1 0.0147

Jet/Plain (2) 1.67x10-4 (10) 18,732 387.2 0.0174 391.2 0.0190

Jet/Microfinned (2) 3.34x10-4 (20)* 376.2 0.0131 380.2 0.0147
8

This report is available at no cost from the National Renewable Energy Laboratory (NREL) at www.nrel.gov/publications.

resistance) is present. The localized cooling from the jets 

eliminates the need for the additional spreading. However, the 

time constant for the heat transfer is decreased, so care should 

be exercised to account for thermal transient effects on the 

instantaneous junction temperature tolerance of the power 

electronic devices.

From previous experiments and modeling, the jetimpingement configurations were expected to have lower 

junction temperatures for the same power level as compared to 

the channel flow configuration. For the full inverter testing, 

the difference (±1.5oC) between thermocouple (±0.8oC) and 

coolant temperatures (±1.2oC) did decrease for the jet

impingement with microfinned surfaces configuration, but 

remained nearly the same for the jet impingement with plain 

surfaces configuration, as seen in Fig. 12.

Fig. 12 Temperature difference between average thermocouple 

and average coolant temperatures.

The location of the thermocouples also affects the 

perceived reduction in temperature between the various 

configurations. They are located near, but not at the hottest 

spot or highest heat flux area. They are also somewhat 

insulated by additional thermal resistance due to thermal 

grease between the modules and cold plate that is used to 

create a more uniform thermal path. The thermocouples are at 

discrete points. It would be ideal to have temperature 

sensors/measurements at each device junction. 

Because the thermocouples are embedded in the middle of 

the module base plate, the difference in temperatures between 

the thermocouples and the coolant represents a thermal 

resistance greater than just the convective resistance and less 

than the total resistance including the stack.

Using the inverter model, we obtained the virtual 

temperature difference between the thermocouple locations 

and the coolant. Using these temperature differences to obtain 

the a comparison of thermal performance between the baseline 

and jet impingement configurations yielded very similar 

reductions to those obtained from comparing the temperature 

differences using the average or maximum device 

temperatures.

When comparing the thermal performance of the jet 

impingement with a plain module configuration to the baseline 

channel flow at full inverter power (all devices powered), we 

observed only a slight improvement in the reduction of the 

temperature difference between the thermocouples and the 

coolant: just -3.4%, 1.2%, 3.6%, and 1.5% for 40, 60, 80, and 

100 kW power levels, respectively (Fig. 13). Adding the 

microfinned enhanced surfaces reduced the temperature 

difference by 6.5%, 14.2%, 16.2%, and 16.5% for 40, 60, 80, 

and 100 kW power levels, respectively. At 100 kW, the 

uncertainty in this reduction was ±6.1%. The reduction in the

temperature difference is directly proportional to the reduction 

in thermal resistance.

Fig. 13 Reduction in temperature difference from baseline 

channel flow to jet impingement configurations.

Using the reduction in temperature difference and pressure 

drop across the heat exchanger, the improvement in the 

coefficient of performance is given in Table 7. Similar to the 

analysis conducted with the modeling, the increase in specific 

power [kW/kg] and power density [kW/L] is also given.

Table 7. Improvement in coefficient of performance, specific 

power, and power density over baseline channel flow.

Power [kW] 40 60 80 100

Coefficient of performance

Jet, Plain -28.5% -0.8% 1.7% -0.5%

Jet, Microfinned 4.8% 14.2% 17.0% 17.4%

Specific Power

Jet, Plain 15.8% 20.1% 22.5% 21.1%

Jet, Microfinned 26.4% 33.6% 35.6% 35.9%

Power Density

Jet, Plain -4.5% -0.9% 1.1% -0.1%

Jet, Microfinned 4.3% 10.2% 11.9% 12.1%

Reliability

Over four months of reliability experimental observations 

are available for this work. After 60 and 120 days, the 

microfinned surface samples were examined for defects and 

thermal conductivity. After an initial break in period, the 

thermal conductivity has remained relatively steady (Fig. 14).

There is no clogging or change in diameter of the nozzles.
9

This report is available at no cost from the National Renewable Energy Laboratory (NREL) at www.nrel.gov/publications.

Fig. 14 Heat transfer coefficients of microfinned enhanced 

surface samples subject to jet impingement.

CONCLUSIONS

A light-weight inverter-level heat exchanger tested at 

inverter power (40-100 kW) using liquid jet impingement and 

microfinned enhanced surfaces showed improved thermal 

performance by up to 17% when compared to a channel flow 

design. Along with a weight reduction of approximately 3 kg, 

the specific power (kW/kg) is improved by up to 36% and 

power density (kW/L) by 12% over the baseline channel flow 

heat exchanger. Jet impingement on plain surfaces yielded a 

slight thermal performance improvement (3%) over the 

baseline. 

Experiments at low power (105 W) demonstrated that the 

thermal resistance is reduced 5% to 13% with plain and 

microfinned enhanced surfaces, respectively, compared to the 

baseline heat exchanger. The base plate-to-coolant 

(convective) resistance was reduced by up to 37% for the jetbased configuration as compared to the baseline, suggesting 

that while improvements to the cooling side reduce the overall 

resistance, efforts to reduce the passive stack resistance may 

also greatly contribute to lowering the overall junction-tocoolant resistance. 

CFD modeling at inverter power, resulting in 2.5 kW of 

heat dissipation predicted a thermal resistance reduction of 9% 

and 32% for the plain and microfinned-enhanced surfaces, 

respectively. The modeling represents an idealized limit if all 

heat from the devices was transferred to the coolant.

Fabrication of the heat exchanger out of light-weight 

plastic may provide additional cost savings associated with 

inexpensive materials and cost-effective manufacturing 

techniques. The cost of surface enhancement must be 

weighed. Effective reliable sealing between the modules and 

the heat exchanger is also vital.

Experiments were conducted to evaluate the reliability of 

the microfinned enhanced surfaces when subjected to nearcontinuous impinging WEG jets. Testing is ongoing, but after 

an initial break-in period, the nozzles and the nickel-plated 

enhanced surfaces have not degraded.

ACKNOWLEDGMENTS

The authors would like to acknowledge the support 

provided by Susan Rogers and Steven Boyd, Technology 

Development Managers for Advanced Power Electronics and 

Electric Motors, Vehicle Technologies Office, U.S. 

Department of Energy Office of Energy Efficiency and 

Renewable Energy. The authors thank Titus Herschberger 

(UQM Technologies Inc.) for full-inverter testing of three 

configurations. The authors would also like to acknowledge 

Jon Lutz (UQM Technologies Inc.) for providing very useful 

information on the UQM inverter, Sy-Jenq Loong (Wolverine 

Tube Inc.) for providing microfinned surface enhancements, 

as well as Ralph Taylor (Delphi) for providing the DBA/DBC 

substrates. 

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